U.S. patent number 10,101,059 [Application Number 12/745,168] was granted by the patent office on 2018-10-16 for thermally driven heat pump for heating and cooling.
This patent grant is currently assigned to The Curators of the University of Missouri, Thermavant Technologies LLC. The grantee listed for this patent is Joseph A. Boswell, Peng Cheng, Hongbin Ma. Invention is credited to Joseph A. Boswell, Peng Cheng, Hongbin Ma.
United States Patent |
10,101,059 |
Ma , et al. |
October 16, 2018 |
Thermally driven heat pump for heating and cooling
Abstract
A thermally driven heat pump includes a low temperature
evaporator for evaporating cooling fluid to remove heat A first
heat exchanger located at an outlet of a converging/diverging
chamber of a first ejector receives a flow of primary fluid vapor
and cooling fluid vapor ejected from the first ejector for
condensing a portion of the cooling fluid vapor An absorber located
in the first heat exchanger absorbs cooling fluid vapor into an
absorbing fluid to reduce the pressure in the first heat exchanger
A second heat exchanger located at an outlet of a
converging/diverging chamber of a second ejector receives primary
fluid vapor and cooling fluid vapor ejected from the second ejector
for condensing the cooling fluid vapor and the primary fluid vapor
A separator in communication with the second ejector, the low
temperature evaporator and the primary fluid evaporator separates
the primary fluid from the cooling fluid.
Inventors: |
Ma; Hongbin (Columbia, MO),
Boswell; Joseph A. (San Francisco, CA), Cheng; Peng
(Coumbia, MO) |
Applicant: |
Name |
City |
State |
Country |
Type |
Ma; Hongbin
Boswell; Joseph A.
Cheng; Peng |
Columbia
San Francisco
Coumbia |
MO
CA
MO |
US
US
US |
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|
Assignee: |
The Curators of the University of
Missouri (Columbia, MO)
Thermavant Technologies LLC (Columbia, MO)
|
Family
ID: |
40678998 |
Appl.
No.: |
12/745,168 |
Filed: |
November 26, 2008 |
PCT
Filed: |
November 26, 2008 |
PCT No.: |
PCT/US2008/084968 |
371(c)(1),(2),(4) Date: |
June 06, 2011 |
PCT
Pub. No.: |
WO2009/070728 |
PCT
Pub. Date: |
June 04, 2009 |
Prior Publication Data
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Document
Identifier |
Publication Date |
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US 20110259039 A1 |
Oct 27, 2011 |
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Related U.S. Patent Documents
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Application
Number |
Filing Date |
Patent Number |
Issue Date |
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60990407 |
Nov 27, 2007 |
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Current U.S.
Class: |
1/1 |
Current CPC
Class: |
F25B
1/08 (20130101); F25B 15/02 (20130101); F25B
2341/0011 (20130101); Y02B 30/62 (20130101); Y02A
30/277 (20180101); Y02A 30/27 (20180101); F25B
2341/0015 (20130101) |
Current International
Class: |
F25B
33/00 (20060101); F25B 1/08 (20060101); F25B
15/02 (20060101) |
Field of
Search: |
;62/101,102,141,495,500 |
References Cited
[Referenced By]
U.S. Patent Documents
Foreign Patent Documents
Other References
NPL--3M Novec 7000 Engineered Fluid, Sep. 2009, pp. 1-6. cited by
examiner .
NPL--Thermodynamics--Property Tables and Charts, Jul. 2009, pp.
1-6. cited by examiner .
Borgmeyer, B. et al., Experimental Investigation of Oscillating
Motions in a Flat Plate Pulsating Heat Pipe, Journal of
Thermophysics and Heat Transfer, vol. 21, No. 2, Apr.-Jun. 2007,
pp. 405-409. cited by applicant .
Cheng, Peng et al., An Investigation of Flat-Plate Oscillating Heat
Pipes, ASME Journal of Electronic Packaging, vol. 132, No. 4
041009, Dec. 2010, 20 pgs. cited by applicant .
Khandekar, S. et al., Thermofluid Dynamic Study of Flat-Plate
Closed-Loop Pulsating Heat Pipes, Microscale Thermophysical
Engineering, 6:303-317 (2002), 15 pgs. cited by applicant .
Thompson, S. M., et al., Experimental Investigation of Miniature
Three-Dimensional Flat-Plate Oscillating Heat Pipe, Journal of Heat
Transfer, vol. 131, Apr. 2009, 10 pages. cited by applicant .
Boswell, Joe, Project entitled Solar Thermal HVAC System Driven by
a High-Efficiency Heat-Pipe Jet Engine, ThermAvant Technologies
LLC, 19 pages, 2007. cited by applicant .
Supplemental European Search Report for EP 08 85 5297 dated Jan. 3,
2014, 11 pages. cited by applicant.
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Primary Examiner: Furdge; Larry
Attorney, Agent or Firm: Senniger Powers LLP
Claims
What is claimed is:
1. A thermally driven heat pump comprising: a low temperature
evaporator for evaporating a cooling fluid to remove heat; a
primary fluid evaporator for evaporating primary fluid by
application of heat; an ejector including a converging/diverging
chamber, nozzle apparatus in fluid communication with the primary
fluid evaporator to receive the primary fluid vapor and to eject
the primary fluid vapor into the converging/diverging chamber, the
low temperature evaporator being in fluid communication with the
converging/diverging chamber so that cooling fluid vapor from the
low temperature evaporator is aspirated into the
converging/diverging chamber; a heat exchanger located at an outlet
of the converging/diverging chamber of the ejector for receiving a
flow of the primary fluid vapor and the cooling fluid vapor ejected
from the ejector for condensing a portion of at least one of the
cooling fluid vapor and the primary fluid vapor; absorption
apparatus located at the outlet of the converging/diverging
chamber, the absorption apparatus including an absorber associated
with the heat exchanger for absorbing the cooling fluid vapor into
an absorbing fluid thereby to reduce a pressure in the heat
exchanger; a generator for separating the cooling fluid from the
absorbing fluid; and a separator in fluid communication with the
heat exchanger, the low temperature evaporator and the primary
fluid evaporator for use in separating the primary fluid from the
cooling fluid for returning to the low temperature evaporator and
the primary fluid evaporator, respectively.
2. A thermally driven heat pump as set forth in claim 1 wherein the
absorption apparatus comprises at least one sprayer arranged to
spray the absorbing fluid into a flow of the cooling fluid and the
primary fluid exiting the outlet of the converging/diverging
chamber.
3. A thermally driven heat pump as set forth in claim 1 in
combination with the absorbing fluid and the primary fluid, wherein
the absorbing fluid is immiscible with the primary fluid.
4. A thermally driven heat pump as set forth in claim 1 further
comprising a return conduit extending between the separator and the
primary fluid evaporator, the return conduit being in thermal
communication with the generator for transferring heat from the
cooling fluid leaving the generator to primary fluid in the return
conduit.
5. A thermally driven heat pump as set forth in claim 1 further
comprising a conduit between the heat exchanger and the generator
for delivering a solution of the absorbing fluid and cooling fluid
absorbed by the absorbing fluid to the generator, the conduit being
in thermal communication with the generator for transferring heat
from the absorbing fluid leaving the generator to the solution of
the absorbing fluid and the cooling fluid entering the
generator.
6. A thermally driven heat pump as set forth in claim 1 wherein the
primary fluid is immiscible with the cooling fluid for separation
by gravity from the cooling fluid in liquid phase and wherein the
primary fluid and cooling fluid have global warming potentials of
less than about 1000, the primary fluid being different from the
cooling fluid and comprising a hydrofluoroether.
7. A thermally driven heat pump as set forth in claim 6 wherein the
ratio of a molecular weight of the primary fluid to a molecular
weight of the cooling fluid is at least about 5.0.
8. A thermally driven heat pump as set forth in claim 6 wherein the
ratio of a heat of vaporization of the cooling fluid to a heat of
vaporization of the primary fluid is at least about 2.0.
9. A thermally driven heat pump as set forth in claim 6 wherein the
primary fluid evaporator includes a liquid reservoir compartment
containing condensed cooling fluid, heat pipes and a wick
separating the liquid reservoir compartment from the heat pipes,
the wick being adapted to draw the cooling fluid through the wick
and into the heat pipes, each heat pipe having an internal surface
structure having microwicks for use in thin film evaporation of
primary fluid.
10. A thermally driven heat pump as set forth in claim 6 wherein
the low temperature evaporator includes a liquid reservoir
compartment containing condensed cooling fluid, heat pipes and a
wick separating the liquid reservoir compartment from the heat
pipes, the wick being adapted to draw the cooling fluid through the
wick and into the heat pipes, each heat pipe having an internal
surface structure having microwicks for use in thin film
evaporation of cooling fluid.
11. A thermally driven heat pump as set forth in claim 6 further
comprising: a second ejector including a converging/diverging
chamber, a nozzle apparatus for ejecting primary fluid vapor into
the second ejector's converging/diverging chamber at high speed,
the heat exchanger being in fluid communication with the
converging/diverging chamber of the second ejector so that cooling
fluid vapor and the primary fluid vapor in the heat exchanger are
aspirated into the converging/diverging chamber of the second
ejector; a second heat exchanger located at an outlet of the
converging/diverging chamber of the second ejector for removing
heat from vapors and the primary fluid vapor to cool the cooling
fluid vapor and the primary fluid vapor to facilitate condensation;
and a conduit to connect the second heat exchanger to the
separator.
12. A thermally driven heat pump as set forth in claim 11 wherein
the primary fluid evaporator constitutes a first primary fluid
evaporator, the heat pump further comprising a second primary fluid
evaporator for evaporation of primary fluid by application of heat,
the second primary fluid evaporator being in fluid communication
with the second ejector.
13. A thermally driven heat pump as set forth in claim 6 further
comprising a return conduit containing the primary fluid connected
to the separator and to the primary fluid evaporator for returning
primary fluid to the primary fluid evaporator, wherein the return
conduit is in thermal communication with the heat exchanger for
pre-heating the primary fluid returning to the primary fluid
evaporator.
14. A thermally driven heat pump as set forth in claim 6 wherein
the converging/diverging chamber is shaped so that the velocity of
the combined flow of the primary fluid and the cooling fluid is
about Mach 1 at a throat of the converging/diverging chamber.
15. A thermally driven heat pump as set forth in claim 1 wherein
the converging/diverging chamber has an inner wall, a centerline
and an inlet adapted for connection to the low temperature
evaporator for aspirating vaporized cooling fluid from the
evaporator, and a rotating section rotatable about the centerline
of the converging/diverging chamber, the rotating section including
projections extending inwardly from the inner wall of the
converging/diverging chamber and into the flow of the primary fluid
and the cooling fluid in the converging/diverging chamber, the
projections being arranged to mix the cooling fluid with the
primary fluid, the projections being shaped to convert some of the
kinetic energy of the flow of primary fluid and cooling fluid to
rotational movement of the rotating section for mixing the primary
fluid and cooling fluid.
Description
BACKGROUND
Refrigeration using a high-pressure jet of steam is typically
referred to as steam jet cooling. In this method, the cooling
system includes a source of steam, an ejector and a closed water
vessel fluidly connected to the ejector and containing a
refrigerant, usually water. In use, the steam is passed through the
ejector to create a partial vacuum in the closed water vessel. Some
of the water in the closed vessel vaporizes at the low pressure of
the partial vacuum and exhausts into a chamber of the ejector. The
vaporized water absorbs heat from the water that remains liquid in
the vessel, thereby cooling the liquid water by evaporative
cooling. The chilled water is pumped through the system to cool
air, and the vaporized water in the ejector is directed to a
condenser, where it condenses into liquid and returned to the
cooling system. Variations of the aforementioned steam jet system
will be known to those of ordinary skill in the art. Systems of
this type could also be used for heating.
In order to be commercially practical, an ejector type heat pump
needs to have a coefficient of performance (COP) of 1.0-or-greater
even when condensing temperatures exceed 100 degrees Fahrenheit. As
is known, COP is the ratio of the cooling power to the input power
required to achieve the cooling. In an ejector type heat pump the
COP is determined by the entrainment ratio (the mass ratio of the
refrigerant fluid to working fluid), and the ratio of the enthalpy
change of the refrigerant fluid to the enthalpy change of the
working fluid. The COP also correlates to the ("lift") ratio of the
pressure of the fluid leaving the ejector to the stagnation
pressure of the refrigerant entering the ejector. The lift ratio,
particularly at high ambient temperatures, requires a high pressure
at the exit of the ejector in order for the vapors to reach their
saturation pressures and to condense. This requires a substantial
amount of heat energy to be supplied to the working fluid, which
increases the enthalpy change of the primary fluid and therefore
reduces the system's efficiency. In fact, a major disadvantage with
the conventional steam jet cooling system is the low coefficient of
performance (COP), which is typically about 0.2 to 0.3. One method
to improve the COP of an ejector system is to choose a refrigerant
fluid (sometimes referred to as the secondary fluid) that is
different from the working fluid (sometimes referred to as the
primary or driving fluid). Such two-fluid jet cooling systems have
achieved COPs of up to 0.5 but have not found commercial
acceptance. Another problem with conventional steam jet cooling
system is the use of non-environmentally friendly fluids as the
working fluid. For example, perfluorocarbon has been used as the
primary fluid because of its high molecular weight and
immiscibility with common refrigerants such as water, acetone,
ammonia, and methanol. However, perfluorocarbons have high global
warming potentials. Moreover, typical ejector systems suffer large
efficiency losses from the shock that accompanies abrupt transition
from supersonic to subsonic flow. Because the losses from a shock
are exponentially related to the pre-shock Mach number, a Mach
number approaching 1.0 or lower can greatly reduce or even
eliminate the shock losses in an ejector system. Still further,
kinetic energy losses can occur as the refrigerant vapor is
accelerated by the working fluid in the ejector.
SUMMARY
In one aspect of the present invention, a thermally driven heat
pump comprising a cooling fluid, a low temperature evaporator for
evaporating the cooling fluid to remove heat, and a primary fluid
immiscible with the cooling fluid for rapid separation by gravity
from the cooling fluid in liquid phase and where primary fluid and
cooling fluids have global warming potentials of less than about
1000. A primary fluid evaporator can be used for evaporating the
primary fluid by application of heat. An ejector includes a
converging/diverging chamber, and nozzle apparatus in fluid
communication with the primary fluid evaporator to receive primary
fluid vapor and to eject the primary fluid vapor into the
converging/diverging chamber at high speed. The low temperature
evaporator is in fluid communication with the converging/diverging
chamber so that cooling fluid vapor from the low temperature
evaporator is aspirated into the converging/diverging chamber. A
first heat exchanger located at an outlet of the
converging/diverging chamber of the ejector can receive a flow of
primary fluid vapor and cooling fluid vapor ejected from the
ejector for removing heat from the cooling fluid vapor and primary
fluid vapor to facilitate condensation of the cooling fluid vapor
and the primary fluid vapor. A separator is in fluid communication
with the heat exchanger, the low temperature evaporator and the
primary fluid evaporator for use in separating the primary fluid
liquid from the cooling fluid liquid, so that cooling fluid can be
returned to the low temperature evaporator and the primary fluid
can be returned to the high temperature evaporator. A return
conduit containing primary fluid is connected to the separator and
to the primary fluid evaporator for returning primary fluid to the
primary fluid evaporator.
In another aspect, an ejector for use in a thermally driven heat
pump that includes an evaporator for evaporating a cooling fluid
generally comprises a converging/diverging chamber having
centerline and an inlet adapted for connection to the evaporator of
the thermally driven heat pump for aspirating vaporized cooling
fluid from the evaporator. A nozzle apparatus is located for
ejecting a high speed vapor flow of a primary fluid into the
converging/diverging chamber. A control controls the nozzle
apparatus to at least one of oscillate and nutate the flow from the
nozzle apparatus generally laterally of the chamber centerline.
In a further aspect of the present invention, an ejector for use in
a thermally driven heat pump including an evaporator for
evaporating a cooling fluid generally comprises a
converging/diverging chamber having an inlet adapted for connection
to the evaporator of the thermally driven heat pump for aspirating
vaporized cooling fluid from the evaporator. A nozzle apparatus is
located for ejecting a high speed vapor flow of a primary fluid
into the converging/diverging chamber. The converging/diverging
chamber includes a rotary mixing section downstream from the nozzle
apparatus adapted to mix primary fluid with the aspirated cooling
fluid.
In yet another aspect, an ejector heat pump system generally
comprises a low temperature evaporator for evaporating a cooling
fluid to remove heat, and a primary fluid evaporator for
evaporating primary fluid by application of heat. A first ejector
includes a converging/diverging chamber and nozzle apparatus in
fluid communication with the primary fluid evaporator to receive
primary fluid vapor and to eject the primary fluid vapor into the
converging/diverging chamber at high speed. The low temperature
evaporator is in fluid communication with the converging/diverging
chamber so that cooling fluid vapor from the low temperature
evaporator is aspirated into the converging/diverging chamber. A
first heat exchanger is located at an outlet of the
converging/diverging chamber of the first ejector for receiving a
flow of primary fluid vapor and cooling fluid vapor ejected from
the first ejector for condensing a portion of at least one of the
cooling fluid vapor and primary fluid vapor. A second ejector
includes a converging/diverging chamber and a nozzle apparatus
adapted to eject a fluid into the converging/diverging chamber at
high speed. The first heat exchanger is in fluid communication with
the converging/diverging chamber of the second ejector so that
cooling fluid vapor and primary fluid vapor in the first heat
exchanger are aspirated into the converging/diverging chamber. A
second heat exchanger is located at an outlet of the
converging/diverging chamber of the second ejector for receiving
primary fluid vapor and cooling fluid vapor ejected from the second
ejector for condensing the cooling fluid vapor and the primary
fluid vapor. A separator is in fluid communication with the second
ejector, the low temperature evaporator and the primary fluid
evaporator for use in separating the primary fluid from the cooling
fluid to be returned to the low temperature evaporator and primary
fluid evaporator, respectively.
In another aspect, a heat pipe jet engine generally comprises a
high velocity nozzle apparatus, and a fluid evaporator fluidly
connected upstream to the nozzle apparatus. The evaporator includes
at least one heat pipe having wicking structure disposed on an
interior surface of the pipe for use in thin film evaporation of
fluid from an evaporator.
In yet another aspect, a method of cooling generally comprises
heating a primary fluid to vaporize the primary fluid, and passing
the vaporized primary fluid through a nozzle of a first ejector
into a high speed flow producing a vacuum pressure adjacent to the
flow. A cooling fluid vapor is aspirated from an evaporator by the
vacuum pressure produced by the nozzle. The primary fluid is
ejected and cooling fluid is entrained into a first heat exchanger.
At least one of the cooling fluid vapor and primary fluid vapor is
partially condensed in the first heat exchanger. The cooling fluid
vapor and primary fluid vapor are aspirated from the first heat
exchanger with a second ejector. The cooling fluid vapor and
primary fluid vapor are ejected from the second ejector into a
second heat exchanger. The primary fluid vapor and cooling fluid
vapor are condensed in the second heat exchanger.
In another aspect, a thermally driven heat pump generally comprises
a low temperature evaporator for evaporating a cooling fluid to
remove heat, and a primary fluid evaporator for evaporating primary
fluid by application of heat. An ejector includes a
converging/diverging chamber and nozzle apparatus in fluid
communication with the primary fluid evaporator to receive primary
fluid vapor and to eject the primary fluid vapor into the
converging/diverging chamber at high speed. The low temperature
evaporator is in fluid communication with the converging/diverging
chamber so that cooling fluid vapor from the low temperature
evaporator is aspirated into the converging/diverging chamber. A
heat exchanger located at an outlet of the converging/diverging
chamber of the ejector receives a flow of primary fluid vapor and
cooling fluid vapor ejected from the ejector for condensing a
portion of at least one of the cooling fluid vapor and primary
fluid vapor. An absorption apparatus includes an absorber located
in the heat exchanger for absorbing cooling fluid vapor into an
absorbing fluid thereby to reduce the pressure in the heat
exchanger. A generator separates the cooling fluid from the
absorbing fluid. A separator is in fluid communication with the
ejector, the low temperature evaporator and the primary fluid
evaporator for use in separating the primary fluid from the cooling
fluid to be returned to the low temperature evaporator and primary
fluid evaporator, respectively.
BRIEF DESCRIPTION OF THE DRAWINGS
FIG. 1 is a schematic of a first embodiment of a thermally driven
heat pump;
FIG. 2A is a schematic of one embodiment of a first working fluid
evaporator of the thermally driven heat pump;
FIG. 2B is a top view schematic of the first working fluid
evaporator;
FIG. 2C is a schematic of another embodiment of the first working
fluid evaporator;
FIG. 2D is a section of the first fluid evaporator of FIG. 2C taken
along the line 2D-2D;
FIG. 3 is a schematic of one embodiment of a second working fluid
evaporator of thermally driven heat pump;
FIG. 4 is a schematic of one embodiment of a first ejector;
FIG. 5A is a schematic of the ejector illustrating flow with two
flow control nozzles open;
FIG. 5B is a schematic of the ejector similar to FIG. 5A except
that one of the flow control nozzles is closed;
FIG. 5C is a schematic of the ejector similar to FIG. 5B except
that the closed nozzle is open and the other of the flow control
nozzles is closed;
FIG. 6A is a schematic of one embodiment of a mixing section of a
converging-diverging chamber of the ejector;
FIG. 6B is a section taken along the line 6B-6B in FIG. 6A.
FIG. 7 is a schematic of a second embodiment of a thermally driven
heat pump; and
FIG. 8 is a schematic of a third embodiment of a thermally driven
heat pump.
Corresponding reference characters indicate corresponding parts
throughout the drawings.
DETAILED DESCRIPTION OF THE DRAWINGS
Referring now to the drawings, and in particular to FIG. 1, a first
embodiment of a thermally driven heat pump is generally indicated
at 10. In use, a first primary or working fluid evaporator 12
supplies a first ejector, generally indicated at 14, with high
pressure, vaporized motive or working fluid (broadly, a primary
fluid) as indicated by arrow A1. The first ejector 14 draws
vaporized refrigerant fluid (broadly, a secondary or cooling fluid)
from a low temperature evaporator 16 (referred to herein as "LTE";
broadly, a second evaporator) into the ejector, as indicated by
arrow A2, where it mixes with the working fluid. An initial heat
exchanger, generally indicated at 18, receives the vaporized
working-refrigerant fluid mixture from an outlet of the first
ejector 14 (indicated at arrow A3). Coolant, such as water or a
glycol-water mixture is circulated through a first coil 19
(broadly, a conduit) of the initial heat exchanger 18. An absorber
20 supplies an absorbent in the initial heat exchanger 18. Instead
of or in addition to the first coil 19, ambient air may be blown
over the heat exchanger 18 to remove heat. The absorbent absorbs
some of the vaporized refrigerant fluid of the vaporized
working-refrigerant fluid mixture exiting the first ejector 14. The
absorbent and the vaporized refrigerant fluid absorbed by the
absorbent are collected in the initial heat exchanger 18, as
indicated by A/R. This absorbent/refrigerant is delivered (e.g.,
pumped) to a generator 21, where the refrigerant is vaporized to
separate the refrigerant from the absorbent. From the generator 21,
the liquid absorbent is pumped back to the absorber 20 in the
initial heat exchanger 18, and the vaporized refrigerant is
delivered into a coil 22 (broadly, a conduit) of a second working
fluid evaporator 23. The heat of the vaporized refrigerant in the
coil 22 vaporizes the working fluid in the second working fluid
evaporator 23, for purposes explained below, and the vaporized
refrigerant condenses in the coil 22 and is delivered to a low
temperature collector or separator, generally indicated at 24. The
separator 24 includes separated layers of liquid refrigerant and
liquid working fluid. From the separator 24, the working fluid is
delivered to the second working fluid evaporator 23, where the
working fluid is vaporized, as explained above. The vaporized
working fluid flows through a second ejector, generally indicated
at 25, to draw the remainder of the vaporized mixture of working
fluid and refrigerant in the initial heat exchanger 18 into the
second ejector. A secondary heat exchanger 26, which functions as a
condenser, is fluidly connected to an outlet of the second ejector
25 and condenses the entrained mixture of refrigerant and working
fluid from the second ejector. From the secondary heat exchanger
26, the condensed mixture of refrigerant and working fluid flows to
the separator 24. The condensed refrigerant in the separator 24 is
delivered to the low temperature evaporator 16, while the condensed
working fluid in the separator is delivered through a second coil
27 (broadly, a second conduit) in the initial heat exchanger 18 to
absorb heat from the vaporized working-refrigerant fluid mixture
before the working fluid is delivered to the first working fluid
evaporator 12.
As will become apparent throughout this discussion of the
embodiments of the invention, the working fluid desirably has a
latent heat of vaporization that is much less than the latent heat
of vaporization of the refrigerant so that the working fluid
vaporizes with a relatively small heat input. In one embodiment,
the ratio of the heat of vaporization of the refrigerant to the
heat of vaporization of the working fluid is at least about 2.0.
Although the illustrated embodiments are concerned with
refrigeration, it will be understood that the thermally driven heat
pump of the present invention has other applications. For instance,
the heat pump may be used to heat, rather than cool a space. Still
further, the heat pump may have application to other apparatus not
specifically purposed for the movement of heat. For example,
evaporation produced by this apparatus could be used of
desalinization or other useful processes.
As will be understood, the optimal design of the heat pump 10 is
dependent on the amount of desired cooling to be obtained by the
system at the low temperature evaporator 16. For purposes of the
below discussion, the desired results and functions of the various
components and aspects of the heat pump will be discussed with the
understanding that the parameters of such components and aspects
are dependent on variables such as, but not limited to, the desired
amount of cooling, the desired amount of work put into the heat
pump at the working fluid evaporators, the thermal characteristics
of the desired working fluid and the desired refrigerant fluid, and
the desired design of the ejector.
The first working fluid evaporator 12 of this embodiment utilizes
hot water from an external source to heat the working fluid in the
first working fluid evaporator 12. It is understood that the heat
source may be other than hot water, such as a gas-driven heater or
biomass-driven heater, or a solar energy collector or process waste
heat source. Other ways of providing heat to the first working
fluid evaporator 12 do not depart from the scope of the present
invention. The temperature of the hot water from the external
source is generally maintained at a temperature of at least about
75.degree. C., and preferably about 120.degree. C., to maintain
relatively high vapor pressure in the first working fluid
evaporator 12. As shown in FIGS. 2A and 2B, the first working fluid
evaporator 12 of the illustrated embodiment comprises an enclosed
chamber 28 having a lower liquid reservoir compartment 30, a heat
transfer compartment 32 disposed above the reservoir chamber and a
vapor compartment 34 disposed above the heat transfer compartment.
The heat transfer compartment 32 is sealed from fluid communication
with the liquid reservoir compartment 30 and vapor compartment 34.
The liquid reservoir compartment 30 is fluidly connected to both
the initial heat exchanger 18 and the separator 24 for receiving
recycled working fluid condensate, indicated at arrow A4. The vapor
compartment 34 is fluidly connected to the ejector 14 for
delivering working fluid vapor, as indicated by arrow A1. A
plurality of heat pipes, each generally indicated at 38, disposed
within the heat transfer compartment 32 have lower open ends in
fluid communication with the liquid reservoir compartment 30 and
upper open ends in fluid communication with the vapor compartment
34. The respective ends of the heat pipes 38 are not in fluid
communication with the heat transfer compartment 32. The heat
transfer compartment includes an inlet 40 for receiving the hot
water, or other fluid, and an outlet 42 for removing the water from
the compartment. The water may be heated by a suitable source of
energy, such as solar, electricity, natural gas or other means at
the source S. The source S may also use waste heat generated in a
separate process. As explained in more detail below when describing
the heat pipes 38, heat from the hot water is absorbed by the
liquid working fluid W.sub.L in the heat pipes to vaporize the
working fluid.
The first working fluid evaporator 12 further includes a wick 44
generally between the liquid reservoir compartment 30 and the heat
pipes 38. The wick 44 is made of a porous material such as a bundle
copper wire filaments (similar to steel wool) that draw liquid from
the liquid reservoir compartment by capillary action upward to the
heat pipes 38. The illustrated first working fluid evaporator 12
comprises an array twenty-one heat pipes 38 having generally
identical structures. It will be understood that other numbers and
configurations of heat pipes may be used within the scope of the
present invention. Each heat pump 38 includes a tubular body 46
having an axial length extending between the open ends of the pump.
The tubular body 46 has an exterior surface and an interior surface
defining an axial passage 48. A plurality of heat fins 50 disposed
along the length of the tubular body 46 extend outward from the
exterior surface of the tubular body, generally transverse to the
longitudinal axis of the body. Microgrooves 52 or other microwicks
are disposed on the interior surface of the tubular body 46. The
microgrooves 52 draw liquid working fluid W.sub.L from the wick 44
into a thin film on the wall of the tubular bodies 46 of the heat
pipes 38.
Heat is transferred from the hot water or other fluid flowing
through the heat transfer compartment 32 to the fins 50 and tubular
bodies 46 of the heat pipes 38 and the heat is further transferred
to the liquid working fluid W.sub.L in the microgrooves 52 in the
heat pipes to produce an efficient, thin film evaporation of the
liquid working fluid W.sub.L inside the heat pipes 38. The thin
film evaporation of the working fluid produces a high heat transfer
coefficient resulting in high vapor pressure at a vapor outlet 56,
which is fluidly connected to the ejector 14. This high vapor
pressure is necessary to produce a high velocity flow rate in the
ejector, as will be explained below. It will be understood that the
amount of vapor pressure for a given input depends upon the type of
working fluid.
The thin film evaporation of the working fluid in the heat pipes 38
of the first working fluid evaporator 12 is highly efficient.
Therefore, the temperature of the hot water in the heat transfer
compartment 32 need be only slightly higher than the boiling point
of the working fluid W.sub.L. For example, the temperature of the
water may need to be only 1.degree. C. warmer than the boiling
temperature of the working fluid W.sub.L the low latent heat of the
working fluid in combination with the efficiency of thin film
evaporation in the heat pipes 38 gives the cooling system 10 a good
coefficient of performance (COP). It is believed the COP of the
cooling system 8 is between 1 and 2. In one example where the
working fluid W.sub.L NOVEC.TM. HFE7300, available from the 3M
Company of St. Paul Minn.), the working liquid can be evaporated at
about 120.degree. C. to produce approximately 24.82 psi (about 171
kPa) of vapor pressure in the vapor compartment 34. In the same
example, the refrigerant can be water. Other fluid pairs are
permissible, but preferably the working fluid and refrigerant are
environmentally friendly fluids with low global warming potential
and low ozone depletion potential. Moreover, the working fluid and
refrigerant are preferably immiscible.
Not only may the working fluid have a relatively high molecular
weight, a relatively low latent heat, and be immiscible with the
refrigerant fluid, it may also have a low Global Warming Potential
(GWP) compared to current refrigerants and working fluids. GWP is a
measure of how much a given mass of greenhouse gas is estimated to
contribute to global warming. It is a relative scale which compares
the gas in question to that of the same mass of carbon dioxide
(whose GWP is by definition 1). A GWP is calculated over a specific
time interval, and a smaller GWP is preferable. For the purposes of
this disclosure the timescale is 100 years. Common refrigerants in
vapor compression cycle cooling (VCC) and refrigeration systems
such as HFC-23 (a hydrofluorocarbon) have high GWPs (14,800 for
HFC-23) while others being used in VCC systems such as HFC-134a (a
hydrofluorocarbon) have lower GWPs (1,430 for HFC-134a). Organic
refrigerants such as water and ammonia have GWPs of 0 and are
therefore highly preferable for reduction of greenhouse gas
emissions related to refrigerant leakage. Prior two-fluid ejector
or jet cooling systems proposed using perfluorocarbons such as
FC-75 as marketed by the 3M Company. FC-75 and other
perfluorocarbons have GWPs of 7,000-plus. In one embodiment, the
working fluid for use in the thermally-driven heat pump will have a
low GWP of approximately 1,000 or lower, and more preferably, 500
or lower. One example of such a working fluid is the aforementioned
NOVEC.TM.HFE7300 (having a GWP of about 200) which can be paired
with a refrigerant such as water to form a low GWP heat pump
system.
It is understood that heat for operating the system may be provided
in other ways besides hot water. For example, in another embodiment
illustrated schematically in FIGS. 2C and 2D, a working fluid
evaporator, indicated generally at 58, is configured to vaporize
the working fluid using solar energy. The working fluid evaporator
58 of this embodiment includes a single heat pump 38 having
microgrooves 52 or other microwicks on its interior surface similar
to the heat pipes in FIG. 2A. The heat pump 38 is able to absorb
solar energy directly and transfer the absorbed solar energy to the
working fluid in the microgrooves 52 to produce thin film
evaporation. Other ways of supplying energy to the heat pump 10 are
within the scope of the present invention.
Referring to FIG. 3, the illustrated low temperature evaporator 16
is similar in structure to the illustrated first working fluid
evaporator 12 in FIG. 2A. The low temperature evaporator 16
comprises an enclosed chamber, indicated generally at 110, having a
lower condensate reservoir compartment 112, a heat transfer
compartment 114 disposed above the reservoir compartment and a
vapor compartment 116 disposed above the heat transfer compartment.
The reservoir compartment 112 includes an inlet port 118 fluidly
connected to separator 24.
Referring still to FIG. 3, a plurality of heat pipes, generally
indicated at 138, are disposed within the heat transfer compartment
114. The heat pipes 138 are similar to the heat pipes 38 of the
first working fluid evaporator 12 in that each pipe includes a
tubular body 140 having an axial length extending between the open
ends of the pump. The tubular body 140 has an exterior surface and
an interior surface defining an axial passage 142. A plurality of
heat fins 144 disposed along the length of the tubular body 140
extend outward from the exterior surface of the tubular body,
generally transverse to a longitudinal axis of the body.
Microgrooves 146 or other microwicks are disposed on the interior
surface of the tubular body 140. Each tubular body 140 has a lower
open end in fluid communication with the liquid reservoir 112 and
an upper open end in fluid communication with the vapor compartment
116. The ends of the heat pipes are not in fluid communication with
the heat transfer compartment 114, which is sealed from both the
reservoir compartment 112 and the vapor compartment 116. A wicking
layer 135 delivers condensate refrigerant fluid RL from the
reservoir 112 to microgrooves 146 of the heat pipes 138. The
wicking layer 135 may be formed in the same way as the wick 44 of
the first working fluid evaporator 12.
The vapor compartment 116 includes a vapor outlet 150 fluidly
connected to a refrigerant vapor inlet 166 (FIG. 4) of the first
ejector 14. As will be explained below, a low pressure created in
the first ejector 14 produces low pressure in the low temperature
evaporator 16, more specifically, the vapor compartment 116, so
that the refrigerant condensate RL within the low temperature
evaporator vaporizes within the heat pipes 138 at a lower
temperature and the refrigerant vapor is drawn into the ejector.
External fluid, such as air or water within a household or
building, flows (e.g., is pumped) into the heat transfer
compartment 114 via an inlet 152 where it is cooled when the
refrigerate condensate RL vaporizes. In other words, when the
refrigerate condensate vaporizes in the heat pipes 138, heat is
absorbed from the fluid flowing through the heat transfer
compartment 114. The amount of heat absorbed depends at least in
part upon the latent heat of the refrigerant and the mass flow of
vaporized refrigerant. The external fluid flows out of the heat
transfer compartment 114 through an outlet 154, where the fluid
(e.g., air) may enter duct work, for example, to cool the household
or building.
The use of thin film evaporation of the refrigerant condensate in
the heat pipes 138 of the low temperature evaporator 16 to absorb
heat from the environment increases the cooling capacity of the
cooling system because the thin film evaporation can significantly
reduce the thermal resistance (i.e., the evaporating heat transfer
coefficient can be significantly increased), which can effectively
cool the chilled water or air for household, building or other
suitable use.
The second working fluid evaporator 23 may have essentially the
same construction as the first working fluid evaporator 12. In
particular, the second working fluid evaporator 23 may take
advantage of thin film evaporation, as described for the first
working fluid evaporator 12. The coil 22 represents the heat
transfer to the working fluid in the second working fluid
evaporator. The actual arrangement can be similar what is shown in
FIG. 2A for the first working fluid evaporator in regard to fluid
from the source S flowing over the heat pipes 38. It will be
appreciated that the second working fluid evaporator may have other
configurations (not shown), including configurations which differ
from the first working fluid evaporator.
Referring now to FIG. 4, in the illustrated embodiment the first
ejector 14 includes a converging-diverging primary nozzle 160,
converging-diverging auxiliary or fluid control nozzles 162, and a
converging-diverging chamber 163. The primary nozzle 160 and the
fluid control nozzles 162 are in fluid communication with the high
pressure vapor flow in the vapor compartment 34 of the first
working fluid evaporator via a shared working fluid inlet 164 at an
inlet end of the first ejector 14. The first ejector 14 also
includes the refrigerant vapor inlet 166 at the first end of the
ejector. The refrigerant vapor inlet 166 is fluidly connected to
the low temperature evaporator 16 for receiving vaporized
refrigerant fluid, as indicated by arrows A2. The angle the flow of
refrigerant fluid makes with a centerline of the first ejector 14
is preferably less than about 45.degree. and more preferably less
than about 15.degree.. Because the angle of entry of the
refrigerant is close to parallel with the direction of flow of the
working fluid exiting the main nozzle 160, kinetic energy losses
associated with changing the direction of flow of the refrigerant
are significantly reduced. Downstream of the nozzles 160, 162 in
the converging/diverging chamber 163 is a mixing section 170, and
downstream of the mixing section is an intermediate section 172 and
a diffuser section 174. In general, the ejector 14 operates under
the Venturi effect. The vaporized working fluid from the first
working fluid evaporator 12 enters the nozzles 160, 162 under high
pressure (e.g., about 172 kPa) and exits the nozzles as a high
velocity jet, thus creating a low pressure at the outlets of the
nozzles corresponding to the location of the mixing section 170.
The vaporized refrigerant fluid in the low temperature evaporator
16 is drawn into the mixing section 170 the chamber 163 via the
refrigerant vapor inlet 166, and the refrigerant vapor is entrained
with the vapor working fluid jet.
One of the key measures of an ejector system's efficiency is its
entrainment ratio, which is the ratio of secondary mass flow (e.g.,
refrigerant fluid) to the primary fluid's mass flow (e.g., working
fluid). Two key determining parameters of entrainment ratio are: a)
the pre-mixing contact area ratio of the two fluids inside the
mixing chamber, and b) the primary fluid's Mach number at nozzle
outlet. In the illustrated embodiment, the control nozzles 162
facilitate mixing of the working and refrigerant fluids more
uniformly and with a larger contact area ratio than with a single
primary nozzle. The sizes of control nozzles 162 are relatively
smaller than the primary nozzle 160, and each control nozzle
includes a valve 176 that can be controlled to open and close.
Referring to FIG. 5A, when both valves 176 of the control nozzles
162 are open, the fluid field of the primary flow is unchanged. It
will be understood that the primary flow would also be unaffected
if valves 176 were closed. Referring to FIGS. 5B and 5C, when one
of the valves 176 is closed and the other valve is open so that
fluid is flowing through the open control nozzle, a low pressure is
created at the open nozzle and fluid flowing from the primary
nozzle 160 is directed toward the fluid flow from the control
nozzle 162. By alternating the valves 176 between on and off
positions, a back and forth oscillating or nutating flow of the
primary flow is produced, which can increase the pre-mixing contact
area of the working fluid and the refrigerant fluid. It will be
understood that there may be more than the two control nozzles 162
illustrated. To produce nutation of the flow jet from the primary
nozzle 160, additional control nozzles 162 are located around the
primary nozzle. The oscillation or nutation is generally with
respect to a centerline of the ejector which corresponds to the
direction of flow of the primary fluid from the primary nozzle 160
when not affected by the control nozzles 162. For example in
operation, the flow from the main nozzle 160 may sweep out a cone
shape. This increase in pre-mixing contact area ratio allows for a
higher entrainment ratio of the ejector 14 and increases the COP of
the entire system. The control nozzles 162 also allow for more
complete mixing in a shorter axial distance and therefore allow for
the ejector 14 to be more compact in size.
After complete mixing within the mixing section 170, the supersonic
mixed vapor has a molecular weight that is based on the mol
fractions of the two immiscible working and refrigerant vapors
being mixed and their respective molecular weights. For example,
the ratio of the molecular weight of the working fluid to the
molecular weight of the refrigerant may be least about 5.0. Because
the refrigerant fluid has a substantially lower molecular weight
than the working fluid, the mixed vapor stream has a resulting
molecular weight that is lower than that of the working fluid flow
prior to mixing. The mixed flow's lower molecular weight means it
has a higher local speed of sound, and therefore a lower Mach
number than the working fluid flow had prior to mixing. The
intermediate section 172 has a shape corresponding to a constant
rate of momentum change (CRMC) curve that helps to minimize shock
losses as the flow mixture enters the diffuser 174 from the
intermediate section 172. The intermediate section 172 has
converging section, a diverging section and a minimum radius
between the sections. The vapor mixture reduces speed to its local
speed of sound (i.e., about Mach 1) or lower as it enters the
minimum radius of the intermediate section 172. The flow then
enters the diverging section without an abrupt transition from a
highly supersonic to subsonic flow and without the resulting shock
losses from such a transition. The shock losses of prior ejectors,
without an intermediate section 172 having the constant rate of
momentum change curve, have resulted in a lower total stagnation
pressure at end of the diffuser section 174. In prior multi-fluid
systems where sonic choking can be avoided if the secondary fluid
has a much lower molecular weight than the primary fluid, the fluid
field has not been controlled to align the secondary flow's
velocity gradient with that of the primary flow in order to
minimize the velocity differences of the two fluids, and therefore
minimize the kinetic energy losses incurred during the mixing
process. The total practical effect of the fluid control method,
the shock loss mitigation, and the kinetic energy loss minimization
is to reduce the size and increase the efficiency of the ejector
system used in the illustrated embodiment.
Another way of mixing the working fluid and refrigerant is
schematically illustrated in FIGS. 6A and 6B. In this
configuration, the mixing section 170 may include a rotating
cylinder 180 with fins 182. The rotating cylinder 180 is preferably
mounted on very low friction bearings, minimizing friction losses.
The fins 182 can have a shape, which can transform the momentum of
the working fluid into a mechanical work, i.e., using a part of
momentum from the working fluid to make the cylinder 180 rotate.
The shapes of the fins 182 can be optimized for various conditions
by experimental investigation and mathematical modeling. As the
working fluid flows through the cylinder 180 with fins 182, the
momentum of the working fluid will make the cylinder rotate. As a
result, the rotating of the cylinder 180 with fins 182 will
effectively mix the working fluid and refrigerant fluid and
increase the effective contact area of the working and refrigerant
fluids and further increase the entrainment ratio of the
refrigerant fluid. It will be understood that the rotating cylinder
180 can be used instead of or in addition to oscillating or
nutating the flow of primary fluid from the primary nozzle 160 with
the control nozzles 162 to mix working fluid flow and refrigerant
fluid flow.
Referring to FIG. 1, the vapor mixture of working/refrigerant fluid
comes into the initial heat exchanger 18 from the outlet of the
first ejector 14 with a total pressure that is determined by the
mol fraction of the working and refrigerant fluids, their
respective initial pressures, temperatures and thermodynamic
properties, the Mach number of the working fluid leaving the
nozzles 160, 162, and the fluid control methods within the ejector
14 to maximize the mixing area and to minimize kinetic energy and
shock losses. The total pressure from the first ejector 14 however,
is limited by the known parameters of an ideal turbine-compressor
which will be known to those versed in the art. In order to
condense the mixed vapors from the first ejector 14, the total
pressure exiting the first ejector must be at least as great as the
sum of the individual saturation pressures of the immiscible vapors
at the given condensing temperature. Ambient conditions determine
the condensing temperature at the first condenser 18 and at high
condensing temperatures (e.g. 40.degree. C.) the first ejector 14
may not be able to produce enough exit pressure to condense both
vapors unless the entrainment ratio were significantly lowered and
therefore the COP of the system were also lowered.
In the illustrated embodiment, a constant entrainment ratio is
allowed even at elevated condensing temperatures by introducing
mixed vapors from the first ejector 14 to the absorber 20 within
the initial heat exchanger 18. The absorber 20 along with the
generator 21 are part of an absorption apparatus. The absorbent
fluid of the absorber 20 is used to absorb some of the refrigerant
vapor from the mixture exiting the first ejector 14. For example, a
Lithium Bromide (LiBr) and water solution with a relatively high
LiBr mol fraction can effectively absorb a significant fraction of
the water vapor refrigerant leaving the first ejector 14. In one
example, the working fluid is generally immiscible with the
absorbent to prevent the need to separate the working fluid from
the absorbent in a generator, which would lower the COP of the
system. In one example, the absorbent may have a low global warming
potential and a low ozone depleting potential in order to minimize
the climate changing effects associated with fluid leaks during
charging, operation, or repair of a refrigeration or cooling
system.
Absorbing some of the refrigerant from the first ejector 14 has at
least three primary benefits. First, absorbing some of the
refrigerant lowers the mol fraction of the refrigerant vapor in the
first heat exchanger 18, and therefore, increases the mol fraction
and partial pressure of the working fluid to allow for condensation
of the working fluid vapor leaving the first ejector 14. Second,
absorbing some of the refrigerant lowers the total load on the
second ejector 25 which has the important effect of lowering the
size and input power needed at the second ejector. Finally, the
absorption generates a heat of solution that can be used to
pre-heat the working fluid in the second coil 27 in the initial
heat exchanger 18 in order to reduce the sensible load on the first
working fluid evaporator 12 and therefore increase the COP of the
system. By theoretical calculation, the sensible heat needed to
raise the temperature of the working fluid from the separator 24 to
the evaporating temperature in the first working fluid evaporator
12 may be as much as 50% of the total input energy needed to power
the system. The heat transferred to the working fluid in second
coil 27 in the initial heat exchanger 18 greatly reduces the amount
of sensible heat needed in the first working fluid evaporator 12 to
evaporate the working fluid.
The following is an example of suitable operating parameters for
the initial heat exchanger 18. In this example, HFE7300 is the
working fluid, water is the refrigerant and a strong LiBr and water
solution is the absorbent. The exit pressure from the first ejector
14 and inside condenser 18 is set to be about 3500 Pa. The
temperature of the first working fluidevaporator 12 is about
120.degree. C. and the temperature within the low temperature
evaporator 16 is about 5.degree. C. The ambient temperatures of
both the coolant entering the first coil 19 in the initial heat
exchanger 18 and the working fluid returning to the first working
fluid evaporator from second coil in the initial heat exchanger 27
are about 40.degree. C. Using theoretical calculations, the mol
ratio of refrigerant flow to working fluid flow exiting the first
ejector 14 and entering the initial heat exchanger 18 is 3:1, i.e.
mol fraction of water in the mixture of HFE7300 and water is 75%
water. The stagnation temperature of the vapor mixture leaving the
first ejector 14 is also calculated to be 83.degree. C., and the
specific heat is nearly the same as that of HFE7300.The counter
flowing HFE7300 within the second coil 27 absorbs both the heat of
solution between the water refrigerant and the LiBr-water absorbent
and the sensible heat from the vapor mixture not absorbed by the
Li-Br absorbent. To the extent that the second coil 27 is unable to
bring the temperature of the mixed vapors and absorbent to the
ambient temperature, the separate coolant flows through the first
coil 19 to bring the temperatures of the vapors and absorbent to
the ambient temperature of 40.degree. C. During the absorption
process an estimated 50% of the water vapor entering the initial
heat exchanger 18 from the first ejector 14 is absorbed by the
strong Lithium Bromide solution. The amount of refrigerant vapor
absorbed may vary with working conditions, and selection range may
also vary within the scope of the present invention, but, it is
believed would typically be in the range of 35% to 65%. The
remaining vapor of HFE7300 and water are drawn out by the second
ejector 25, and the pressure in the primary heat exchanger 18 would
be kept at 3500 Pa by the constant suction pressure into the second
ejector. Under these conditions none of the remaining HFE7300 or
water vapor would be condensed in the secondary heat exchanger 26.
The only liquid collected in the initial heat exchanger 18 (e.g.,
at the bottom of the exchanger) would be the LiBr-water mixture
that results from the absorption process described above. It should
be noted that for different combinations of working fluids, the
operating temperatures and pressures may be different from that
described above. Moreover, some of the working fluid and
refrigerant may be condensed in the initial heat exchanger 18.
In order for the absorbent to efficiently absorb the refrigerant
vapor, the absorption process must be efficient and the system must
be compact. In the illustrated embodiment, the absorber 20 includes
one or more miniature or compact jets 186 impinging mixing process
or spray process. The spraying or jet impinging of LiBr directly
into the mixture of the working fluid and refrigerant fluid flows
will result in an increase of the effective contact area between
the working/refrigerant fluid vapor and absorbent, which can
increase the absorption rate of refrigerant vapor or working fluid
and which can reduce the size and the cost of the absorption system
employed.
In the illustrated embodiment, the absorption/refrigeration cycle
between the initial heat exchanger 18 and the generator 21 is an
absorption sub-cycle. The temperature of generator 21 is set to be
a relatively high temperature, for example, 120.degree. C., and the
pressure in the generator may be slightly greater than 200 kPa. The
LiBr weak solution (i.e., A/R mixture) exits the initial heat
exchanger 18 with low pressure and is pumped into the generator 21
by a pump 190. In the generator 21, the absorbent-refrigerant
mixture is heated to separate the mixture into a refrigerant vapor
and a LiBr strong liquid solution. The thermal energy used to heat
the absorbent-refrigerant mixture in the generator 21 may originate
from the same source used for the first working fluid evaporator 12
(e.g., gas- or biomass-fired heater, a solar energy collector, or a
process waste heat source). A pressure drop between the generator
21 and the absorber 20 in the initial heat exchanger 18 causes the
absorbent in the generator to flow to the absorber 20. A valve 198
is used to control the mass flow rate of the absorbent to the
absorber. The absorbent may be delivered from the generator 21 to
the absorber 20 in other ways, including a pump. A heat exchanger
200 transfers heat between the strong absorbent (approx 120.degree.
C.) leaving generator 21 and the weak absorbent (approx 40.degree.
C.) entering the generator from the initial heat exchanger 18 to
lower the temperature of the strong absorbent entering the initial
heat exchanger and to raise the temperature of the weak absorbent
entering the generator. This heat exchanger 200 improves the system
efficiency by raising the amount of refrigerant absorbed in the
initial heat exchanger 18, lowering the heat removal energy in the
first coil 19 of the initial heat exchanger 18, and lowering the
sensible heat required in the generator 21.
Refrigerant vapor leaving generator 21 flows through the coil 22 in
the second working fluid evaporator 23 and is condensed into
liquid. Working fluid from the separator 24 is pumped by a pump 214
into to the second working fluid evaporator 23, whereby the heat
from the vaporized refrigerant in the coil 22 is transferred to the
working fluid to vaporize the working fluid. In this way much of
the energy penalty associated with use of an absorption system is
recovered in the second working fluid evaporator 23. Preferably,
the absorbent apparatus is configured so that the heat released by
the condensing refrigerant in the coil 22 substantially equals the
amount of heat necessary to vaporize the working fluid in the
second evaporator 23 to produce the necessary mass flow through the
second ejector 25 to aspirate the vaporized working-refrigerant
fluid in the first heat exchanger 18. This may affect the amount of
refrigerant vapor selected to be absorbed by the absorber 20.
Preferably, only enough refrigerant vapor is absorbed to provide
the heat when condensed in coil 22 needed to evaporate the working
fluid in the second working fluid evaporator 23. However, other
factors such as the ambient temperature (i.e., temperature of the
environmental cooling source), the desired low temperature
evaporator temperature, the heat source temperature in the first
primary fluid evaporator 12 and the choice of fluids also affect
the amount of refrigerant vapor that is absorbed.
After the condensed refrigerant exits the second working fluid
evaporator 23, the refrigerant flows through a heat exchanger 202
where working fluid being pumped from the separator 24 absorbs heat
from the condensed refrigerant before the working fluid enters the
second working fluid evaporator. In effect, the heat exchanger 202
critically raises the temperature of the working fluid entering
second working fluid evaporator 23 to reduce the sensible heat
needed in the second working fluid evaporator to vaporize the
working fluid and increase the COP of the system. A valve 204
controls the mass flow rate of condensed refrigerant entering the
separator 24.
The vaporized working fluid in the second working fluid evaporator
23 flows under high pressure to the second ejector 25. The
vaporized working fluid flowing through the second ejector 25 draws
the working-refrigerant fluid mixture from the initial heat
exchanger 18 into the second ejector. The second ejector in the
illustrated embodiment includes a mono-nozzle 205 and a
conventional converging-diverging chamber 206. It is contemplated
that the second ejector 25 include a multi-nozzle design similar to
the first ejector 14, and an intermediate section of the
converging-diverging chamber may have a constant rate of momentum
change curve, similar to the first ejector. The second ejector 25
provides at least two functions. First, the second ejector 25
maintains a constant pressure in the fluidly connected initial heat
exchanger 18. Second, the second ejector 25 compresses the mixed
working and refrigerant vapors from the initial heat exchanger 18
to a total pressure exiting the second ejector that is at least as
great as the sum of the individual saturation pressures of the
working and refrigerant vapors. At this critical total pressure,
all of the working and refrigerant vapors exiting the second
ejector 25 can be condensed in the secondary heat exchanger 26. It
is believed that prior single-stage, immiscible fluid pair ejectors
cannot reach this critical pressure without significantly raising
the initial temperature and pressure of the working fluid
evaporator which has the effect of lowering the total COP of the
system and raising the operating costs of the system by a higher
quality, higher temperature energy source.
The secondary heat exchanger 26 is simpler than the initial heat
exchanger 18, although the secondary heat exchanger may have a
design similar to the initial heat exchanger without departing from
the scope of the present invention. In that event, for example,
some of the heat in the secondary heat exchanger 26 may be used to
further pre-heat the working fluid on its way back to the first
working fluid evaporator 12. Coolant, such as water or a
glycol-water mixture, flows through a coil 210 (broadly, a conduit)
in the secondary heat exchanger 26 to absorb heat from the mixed
working and refrigerant vapors exiting the second ejector 25. The
loss of heat to the coolant flowing through the coil 210 condenses
the working fluid and the refrigerant fluid in the secondary heat
exchanger 26. The condensed working fluid and the condensed
refrigerant flow to the separator. All of the working fluid (e.g.,
HFE7300) from the first working fluid evaporator 12 and refrigerant
(e.g., water) from the low temperature evaporator 16 are in liquid
phase when the fluids enter the separator 24. It should be noted
that some of the refrigerant goes through the absorption sub-cycle
but would come back to separator 24 finally. In one example, the
working fluid and the refrigerant are immiscible so that the
condensed working fluid and the condensed refrigerant separate in
the separator 24 by gravity. As previously stated for the
embodiments described herein, HFE7300 can be used as the working
fluid and water can be used as the refrigerant. Other fluid pairs
are possible, although it is preferred that other fluid pairs meet
the following specifications: 1) the fluids have a large difference
of the molecular weight; 2) the fluids have a large difference of
latent heat; 3) the one fluid with higher latent heat is easily
absorbed by the absorbent in the absorption sub-cycle; and 4) the
two fluids are immiscible. In the separator 24, the working fluid
with higher density and lower latent heat, such as HFE7300, would
be located at a bottom of the separator because of its larger
density. The refrigerant, with lighter molecular weight and higher
latent heat, such as water, floats on top of the working fluid.
The refrigerant in the separator 24 flows from the separator to the
low temperature evaporator 16. A valve 212 regulates the flow of
the liquid refrigerant R from the separator 24 to the low
temperature evaporator 16. A pump or other device may be used to
deliver the refrigerant to the low temperature evaporator 16
without departing from the scope of the invention. The working
fluid W in the separator 24 is pumped separately to the first
working fluid evaporator 12 and to the second working fluid
evaporator 23 by respective pumps 214, 216. The pumps, 214, 216 may
be of any suitable type, and may be thermally-driven pumps which
would allow the entire heat pump 10 to operate without electricity.
The pump 216 pumps the working fluid through the second coil 27 of
the initial heat exchanger before the working fluid is delivered to
the first working fluid evaporator 12. As stated above, the working
fluid absorbs heat from the vaporized working-refrigerant mixture
in the initial heat exchanger 18 to reduce the amount of heat that
the working fluid needs to absorb in the first working fluid
evaporator 12 to vaporize the fluid.
Referring now to FIG. 7, a second embodiment of a thermally driven
heat pump is generally indicated at 10'. This embodiment is similar
to the first embodiment, with like components indicated by
corresponding reference numerals plus a single prime. The second
embodiment of the thermally driven heat pump 10' has a simpler
structure that the first embodiment because the second embodiment
does not include an absorption apparatus. A single working fluid
evaporator 12', similar to the first working fluid evaporator 12 of
the first embodiment, supplies vapor working fluid to both a first
ejector 14' and a second ejector 25'. Using the vaporized working
fluid, the first ejector 14' draws in vaporized refrigerant fluid
from a low temperature evaporator 16'. The first ejector 14' may be
similar in structure to the first ejector 14 in the first
embodiment. The vaporized working-refrigerant mixture is entrained
in the first ejector 14' and flows into an initial heat exchanger
18'. At the initial heat exchanger 18', coolant is pumped through
first coils 19' and condensed working fluid is pumped through
second coils 27' to remove heat from the vapor mixture. A portion
of the vapor mixture may condense depending on the components
partial pressure and the total pressure in the primary heat
exchanger. The portion of the mixture that is condensed is
delivered to a separator 24' using a pump 190'. The condensed
working fluid and the condensed refrigerant separate in the
separator 24'. The portion of the mixture that remains vapor is
drawn into the second ejector 25'. In the second ejector, the
working fluid vapor and the refrigerant vapor are entrained and
flow into a second heat exchanger 26' where the vapor mixture is
substantially completely condensed. The condensed working fluid and
refrigerant are delivered to the separator 24' where the fluids are
separated. From the separator 24', the condensed refrigerant flows
to the low temperature evaporator 16'. A valve 212' controls the
flow of the condensed refrigerant to the low temperature evaporator
16'. The condensed working fluid in the separator 24' flows through
the second coil 27' in the initial heat exchanger 18' before the
fluid is delivered to the working fluid evaporator 12' using a pump
216'.
Referring to FIG. 8, a third embodiment of a thermally-driven heat
pump is generally indicated at 10''. This embodiment is similar to
the first embodiment, with like components indicated by
corresponding reference numerals plus a double prime. The third
embodiment of the thermally driven heat pump 10'' is different than
the first embodiment in that the third embodiment includes a single
working fluid evaporator 23'', similar to the second working fluid
evaporator 23 in the first embodiment, and does not include a
second ejector. The working fluid evaporator 23'' supplies vapor
working fluid to an ejector 14'', which may be similar to the
ejector 14 in the first embodiment. The ejector 14'' draws in
vaporized refrigerant fluid from a low temperature evaporator 16''.
The vaporized working fluid and the vaporized refrigerant fluid are
entrained in the ejector 14'' and flow into a heat exchanger 18'',
which is similar to the heat exchanger 18 in the first embodiment.
An absorber 20'' in the heat exchanger 18'' releases absorbent that
absorbs some of the refrigerant vapor. Coolant flowing through
first coils 19'' and working fluid flowing through second coils
27'' remove heat from the working-refrigerant vapor mixture. The
working-refrigerant vapor mixture condenses in the heat exchanger
18'', and the condensed fluids flow into a separator 24'' where
they separate into a layer of absorbent-refrigerant liquid and a
layer of working fluid liquid. From the separator 24'', the working
fluid liquid is pumped, via pump 216'', through the second coil
27'' of the heat exchanger 18'' to the working fluid evaporator
23''. The absorbent-refrigerant liquid is pumped, via pump 214'' to
a generator 21'' where the refrigerant is separated from the
absorbent by vaporizing the refrigerant. From the generator 21'',
the absorbent is delivered to the absorber 20'' at the heat
exchanger 18''. The vaporized refrigerant is delivered through a
coil 22'' in the working fluid evaporator 23'' where heat from the
refrigerant is absorbed by the refrigerant to vaporize the
refrigerant. The refrigerant condenses in the coil 22'' and flows
through a heat exchanger 202'' to transfer additional heat to the
working fluid before the working fluid enters the working fluid
evaporator 23''. A valve 204'' controls the flow of refrigerant
from the coil 22''. From the heat exchanger 202'', the refrigerant
flows to the low temperature evaporator 16''.
Having described the invention in detail, it will be apparent that
modifications and variations are possible without departing from
the scope of the invention defined in the appended claims. It is
envisioned that the cooling system described herein can be used in
numerous situations where cooling and/or heating is needed. The
compact size of the system makes it applicable to automobiles.
There, waste heat from the engine cooling circuit and/or exhaust
gases can be used to drive the cooling system described above.
When introducing elements of the present invention or the preferred
embodiment(s) thereof, the articles "a", "an", "the" and "said" are
intended to mean that there are one or more of the elements. The
terms "comprising", "including" and "having" are intended to be
inclusive and mean that there may be additional elements other than
the listed elements.
As various changes could be made in the above constructions,
products, and methods without departing from the scope of the
invention, it is intended that all matter contained in the above
description and shown in the accompanying drawings shall be
interpreted as illustrative and not in a limiting sense.
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